The invention relates to hydraulic servoactuator devices, and more particularly to such devices that are suitable for use in automobile hydraulic rack and pinion power steering systems, thrust vector control systems for missiles, robotics devices, and the like.
A hydraulic servoactuator is an accurate and high bandwidth position control device for heavy loads. A generic hydraulic servoactuator consists, in general, of a hydraulic power system circulating fluid in a closed loop hydraulic line, a flow control valve transferring the fluid flow rate in both directions between the hydraulic line and the actuator, and an actuator with a power piston that moves the load to the commanded position in response to fluid power. The flow control valve responds to an error signal indicating the relationship between the command and instantaneous positions of the load, and diverts the pump flow rate to the power piston in such a way as to reduce the error signal towards zero in accordance with a nonlinear valve profile curve. The shape of the curve is application specific, and its slope is the valve gain which is critical in determining the bandwidth of a hydraulic servoactuator.
The valve profile curve describes the physical and operational characteristics of a hydraulic servoactuator. The steeper the slope of the curve at a specific error signal the faster will be the rise in pressure on the power piston. Such a hydraulic servoactuator will have a high bandwidth in load positioning. However, as described below, it may cause limit cycles. At time t=ts (where ts is the steering command starting time) a command is applied which yields an instantaneous error signal denoted by e=em (where e is the error, or relative rotary spool valve angle, and em is the mean error, both in degrees). This corresponds to a differential pressure signal denoted by dppp=dpm on the power piston (where dppp is the differential power piston pressure and dpm is the mean differential power piston pressure, both in psi). Referring to FIG. 1, the pressure dpm is equal to P1-P2 and results from the fluid flow rates difference, Q1-Q2. As the power piston position, denoted by xp (where xp is the power piston displacement from central position, in inches), changes under the action of P1-P2, an oscillatory error signal develops and is superimposed on e=em. In the exemplary embodiment of FIG. 1, if the direction of the flow rates (Q1 and Q2) is reversed, then the direction of xp and xpr is also reversed. This oscillatory error signal is induced by the oscillatory motion of the power piston resulting from the pump drive torque oscillations represented by disturbance 1 and load oscillations represented by disturbance 2. A small oscillatory error signal causes a large oscillatory pressure with the opposite polarity on the power piston. This is a salient feature of the hydraulic servoactuator operation where the flow control valve generates a force that opposes the power piston motion caused by a disturbance. The higher the flow control valve gain, the higher the amplitude of the oscillatory component of the differential pressure dppp on the power piston. The force resulting from dppp usually acts opposite to the oscillatory component of the power piston motion. This interaction between the nonlinear flow control valve operation and the oscillatory power piston motion, caused by the disturbances described above, results in pressure oscillations observed in hydraulic servoactuators such as automobile anti-lock brake and steering systems, missile thrust vector control systems and robotic devices.
If pressure oscillations in the two fluid chambers on both sides of the power piston have unequal amplitudes and the same frequencies, then an unbalanced or net oscillatory pressure wave results and, as will be understood by those skilled in the art, an oscillatory net reaction force is therefore generated at the power actuator support. If the frequency of such oscillatory reaction force approximates the structural frequencies, then vibrations will occur. If pressure oscillations in the two fluid chambers on both sides of the power piston have equal amplitudes and the same frequencies, then an unbalanced oscillatory pressure wave will not occur, an oscillatory reaction force will not be generated at the support, and vibrations will not occur.
In an automobile power steering system, the error signal is generated by a relative angle between the valve spool, connected to the steering wheel, and the valve body, rotating with the pinion, which is geared to the rack connected rigidly to the power piston. Therefore, in this application, the steering wheel is the position command device, the pinion is the position feedback sensor, and the rotary spool valve is the combined error detector and flow control valve. In a missile thrust vector control system, the error signal is generated by a relative angle between the reference nozzle tilt angle, generated by the autopilot or the flight computer, and instantaneous nozzle tilt angle, measured by a linear variable differential transformer (LVDT) sensor connected electrically to the power piston shaft. Therefore, in this case the autopilot is the position command device, the LVDT sensor is the position feedback sensor, and the servo valve is the combined error detector and flow control valve. If the large oscillatory pressure and the oscillatory power piston motion have the same polarity or zero phase angle, then it is probable that unstable pressure oscillations will occur and the vibration intensity will increase.
During automobile parking maneuvers, involving large steering angles at slow rates, a vibration type known as steering shudder with a moaning sound occurs. Engine torque oscillations, resulting in pump flow rate oscillations, and diversion of the pump flow to the power piston, develops into an unbalanced oscillatory force on the power piston.
The oscillatory motion on the power piston will demand an opposing cyclic hydraulic force from the rotary spool valve. Automobile rack and pinion power steering systems are servoactuators in which the rotary spool valve is the controller which determines an indicated hydraulic power demand in response to the relative angle between steering wheel and pinion rotation angles. Any oscillatory motion on the power piston, which is directly connected to the rack and pinion, will be fed to the rotary spool valve by the pinion acting as a feedback sensor. Since the rotary spool valve has a high power amplification, or gain, at a large relative angle, and since it always generates a large control force opposing the disturbance force on the power piston, oscillations in the forms of limit cycles will be set up across the power piston.
This unbalanced oscillatory force must be balanced by an equal and opposing reaction force at the steering gear housing frame supports. If this reaction force has a high amplitude and frequency to excite the structural frequencies, then vibrations will occur accompanied by moaning sound. This condition is called steering shudder. Since the power piston is directly connected to the rack, steering wheel vibrations will also occur if the steering shaft structural frequencies are excited by the unbalanced oscillatory force.
During city driving on bumpy roads, and highway driving, rack and pinion oscillations, caused by unequal forces from front tires and brakes, will result in a vibration type known as steering shimmy. If the unbalanced oscillatory force acting on the power piston, caused by rack and pinion oscillations, excites the structural frequencies in the steering shaft, then steering wheel vibrations will occur. In this type vibration, if the oscillatory reaction force at the steering gear housing frame supports excites the structural frequencies, then vibrations and a whining sound will occur.
Prior art has addressed the problems of vibration by means of dampening devices which isolate the vibration from the chassis or steering wheel. Exemplary prior art systems are described in U.S. Pat. No. 4,588,198 to Kanazawa, et al., U.S. Pat. No. 5,157,897 to Satoh et al., and U.S. Pat. No. 5,392,882 to Mackovjak, et al.
Some of these methods involve passive and/or active damping control devices. Most commercially used power steering system engineers apply methods involving: 1) using tuned and longer hoses in hydraulic lines, 2) reducing rotary spool valve gain, and 3) using reinforced elastomer dampers at joints in the steering shaft, tie rods and steering gear housing frame supports.
The first method described above aims to shift the fluid line frequencies away from structural frequencies. This reduces the amplitude of the unbalanced oscillatory force on the power piston, which generates smaller oscillatory reaction force at the steering gear housing frame supports and steering wheel. The second method aims to reduce the amplitude of reaction forces and reducing the unbalanced oscillatory force on the power piston. However, this method also reduces the mean value of the unbalanced force on the power piston which decreases hydraulic power assist for the driver during steering demands. The third method provides passive damping to reduce the amplitude of structural vibrations.
It is known that a 20% increase in fluid line length will demand a 20% increase in pump power for the same design fluid flow rate. It is also known that a tuner reduces the flow rate by 10% because of an increase in friction, while increasing the fluid line length, and hence requires more power from the pump. It is estimated that the combined power consumption of the above three methods is about 30% of the pump power rating.
The prior art has not recognized that the most efficient way of cancelling or reducing structural vibrations is to connect in parallel the two chambers around the power piston of the servoactuator through a pair of fluid delivery lines to the two corresponding chambers of a vibration suppressor, which chambers are separated by a piston which is supported by a pair of springs and tuned to the frequency of the unbalanced oscillatory pressure which causes the vibration. The tuning frequency is dependent on the fluid inductance of the delivery line and spring capacitance. The present invention provides an advantage over the prior art in that it positively reduces and cancels the unwanted vibrations, basically without the need of longer hoses and tubing, tuned hoses, reinforced elastomer dampers at joints in steering shafts, tie rods, and steering gear housing frame supports, or rotary spool valve with modified gain profile to allow leakage, although it is possible to use such prior art systems in combination with the present invention.
The novel features of the invention are set forth with particularity in the appended claims. The invention will be best understood from the following description when read in conjunction with accompanying drawings.